In the field of refrigeration and chillers, the evaporator heat exchanger is a large structure, containing a plurality of parallel tubes, within a larger vessel comprising a shell, through which refrigerant flows, absorbing heat and evaporating. Outside the tubes, an aqueous medium, such as brine, circulates and is cooled, which is then pumped to the process region to be cooled. Such an evaporator may hold hundreds or thousands of gallons of aqueous medium with an even larger circulating volume.
Mechanical refrigeration systems are well known. Their applications include refrigeration, heat pumps, and air conditioners used both in vehicles and in buildings. The vast majority of mechanical refrigeration systems operate according to similar, well known principles, employing a closed-loop fluid circuit through which refrigerant flows, with a source of mechanical energy, typically a compressor, providing the motive forces.
Typical refrigerants are substances that have a boiling point below the desired cooling temperature, and therefore absorb heat from the environment while evaporating under operational conditions. Thus, the environment is cooled, while heat is transferred to another location where the latent heat of vaporization is shed. Refrigerants thus absorb heat via evaporation from one area and reject it via condensation into another area. In many types of systems, a desirable refrigerant provides an evaporator pressure as high as possible and, simultaneously, a condenser pressure as low as possible. High evaporator pressures imply high vapor densities, and thus a greater system heat transfer capacity for a given compressor. However, the efficiency at the higher pressures is lower, especially as the condenser pressure approaches the critical pressure of the refrigerant. It has generally been found that the maximum efficiency of a theoretical vapor compression cycle is achieved by fluids with low vapor heat capacity, associated with fluids with simple molecular structure and low molecular weight.
Refrigerants must satisfy a number of other requirements as best as possible including: compatibility with compressor lubricants and the materials of construction of refrigerating equipment, toxicity, environmental effects, cost availability, and safety.
The fluid refrigerants commonly used today typically include halogenated and partially halogenated alkanes, including chlorofluorocarbons (CFCs), hydrochlorofluorocarbons (HFCFs), and less commonly hydrofluorocarbons (HFCs) and perfluorocarbons (PFCs). A number of other refrigerants are known, including propane and fluorocarbon ethers. Some common refrigerants are identified as R11, R12, R22, R500, and R502, each refrigerant having characteristics that make them suitable for different types of applications.
A refrigeration system typically includes a compressor, which compresses gaseous refrigerant to a relatively high pressure, while simultaneously heating the gas, a condenser, which sheds the heat from the compressed gas, allowing it to condense into a liquid phase, and an evaporator, in which the liquefied refrigerant is again vaporized, withdrawing the heat of vaporization from a process. The compressor therefore provides the motive force for active heat pumping from the evaporator to the condenser. The compressor typically requires a lubricant, in order to provide extended life and permit operation with close mechanical tolerances. Normally, the gaseous refrigerant and liquid lubricant are separated by gravity, so that the condenser remains relatively oil free. However, over time, lubricating oil migrates out of the compressor and its lubricating oil recycling system, into the condenser. Once in the condenser, the lubricating oil becomes mixed with the liquefied refrigerant and is carried to the evaporator. Since the evaporator evaporates the refrigerant, the lubricating oil accumulates at the bottom of the evaporator. The oil in the evaporator tends to bubble, and forms a film on the walls of the evaporator tubes. In some cases, such as fin tube evaporators, a small amount of oil enhances heat transfer and is therefore beneficial. In other cases, such as nucleation boiling evaporator tubes, the presence of oil, for example over 1%, results in reduced heat transfer. See, Schlager, L. M., Pate, M. B., and Berges, A. E., “A Comparison of 150 and 300 SUS Oil Effects on Refrigerant Evaporation and Condensation in a Smooth Tube and Micro-fin Tube”, ASHRAE Trans. 1989, 95(1):387-97; Thome, J. R., “Comprehensive Thermodynamic Approach to Modelling Refrigerant-Lubricating Oil Mixtures”, Intl. J. HVAC&R Research (ASHRAE) 1995, 110-126; Poz, M. Y., “Heat Exchanger Analysis for Nonazeotropic Refrigerant Mixtures”, ASHRAE Trans. 1994, 100(1)727-735 (Paper No. 95-5-1).
Several mechanisms are available seeking to control lubricating oil buildup in the evaporator. One mechanism provides a shunt for a portion of mixed liquid refrigerant and oil in the evaporator to the compressor, wherein the oil is subject to the normal recycling mechanisms. This shunt, however, may be inefficient and is difficult to control. Further, it is difficult to achieve and maintain low oil concentrations using this method.
It is also known to periodically purge the system, recycling the refrigerant with purified refrigerant and cleaning the system. This technique, however, generally permits rather large variance in system efficiency or relatively high maintenance costs. Further, this technique generally does not acknowledge that there is an optimum level of oil in the evaporator and, for example, the condenser. Thus, typical maintenance seeks to produce a “clean” system, subject to incremental changes after servicing.
It is thus known that the buildup of large quantities of refrigerant oil within an evaporator, which passes in small amounts from the compressor to the condenser as a gas, and which leaves the condenser and passes to the evaporator as a liquid, will reduce efficiency of the system, and further, it is known to provide in-line devices which continuously remove refrigerant oil from the refrigerant entering the evaporator. These devices include so-called oil eductors.
The inefficiency of these continuous removal devices is typically as a result of the bypassing of the evaporator by a portion of the refrigerant, and potentially a heat source to vaporize or partially distill the refrigerant to separate the oil. Therefore, only a small proportion of the refrigerant leaving the condenser may be subjected to this process, resulting in poor control of oil level in the evaporator and efficiency loss.
It is also known to reclaim and recycle refrigerant from a refrigeration system to separate oil and provide clean refrigerant. This process is typically performed manually and requires system shutdown.
Systems are available for measuring the efficiency of a chiller, i.e., a refrigeration system which cools water or a water solution, such as brine. In these systems, the efficiency is calculated based on Watt-hours of energy consumed (Volts×Amps×hours) per cooling unit, typically tons or British Thermal Unit (BTU) (the amount of energy required to change the temperature of one British ton of water 1° C.
Thus, a minimal measurement requires a clock, voltmeter, ammeter, and thermometers and flowmeters for the inlet and outlet water. Typically, further instruments are provided, including a chiller water pressure gage, gages for the pressure and temperature of evaporator and condenser. A data acquisition system is also typically provided to calculate the efficiency, in BTU/kWH.
The art, however, does not provide systems intended to measure the operating efficiency of commercial chillers, while permitting optimization of the system.
It is known that the charge conditions of a chiller may have a substantial effect on both system capacity and system operating efficiency. Simply, the level of refrigerant charge in a chiller condenser directly relates to the cooling capacity of the system, all other things being equal. Thus, in order to handle a larger heat load, a greater quantity of refrigerant is required. However, by providing this large refrigerant charge, the operating efficiency of the system at reduced loads is reduced, thus requiring more energy for the same BTU cooling. Bailey, Margaret B., “System Performance Characteristics of a Helical Rotary Screw Air-Cooled Chiller Operating Over a Range of Refrigerant Charge Conditions”, ASHRAE Trans. 1998 104(2), expressly incorporated herein by reference. Therefore, by correctly selecting the “size” (e.g., cooling capacity) of the chiller, efficiency is enhanced. However, typically the chiller capacity is determined by the maximum expected design load, and thus for any given design load, the quantity of refrigerant charge is dictated. Therefore, in order to achieve improved system efficiency, a technique of modulation recruitment is employed, in which one or more of a plurality of subsystems are selectively activated depending on the load, to allow efficient design of each subsystem while permitting a high overall system load capacity with all subsystems operational. See, Trane “Engineer's Newsletter” December 1996, 25(5):1-5. Another known technique seeks to alter the rotational speed of the compressor. See, U.S. Pat. No. 5,651,264, expressly incorporated herein by reference.
Chiller efficiency generally increases with chiller load. Thus, an optimal system seeks to operate system near its rated design. Higher refrigerant charge level, however, results in deceased efficiency. Further, chiller load capacity sets a limit on the minimum refrigerant charge level. Therefore, it is seen that there exists an optimum refrigerant charge level for maximum efficiency.
Chiller efficiency depends on several factors, including subcooling temperature and condensing pressure, which, in turn, depend on the level of refrigerant charge, nominal chiller load, and the outdoor air temperature. First, subcooling within the thermodynamic cycle will be examined. FIG. 6A shows a vapor compression cycle schematic and FIG. 6B shows an actual temperature-entropy diagram, wherein the dashed line indicates an ideal cycle. Upon exiting the compressor at state 2, as indicated in FIG. 6A, a high-pressure mixture of hot gas and oil passes through an oil separator before entering the tubes of the remote air-cooled condenser where the refrigerant rejects heat (Qh) to moving air by forced convection. In the last several rows of condenser coils, the high-pressure saturated liquid refrigerant should be subcooled, e.g., 10 F. to 20 F (5.6 C to 11.1 C), according to manufacturer's recommendations, as shown by state 3 in FIG. 6B. This level of subcooling allows the device following the condenser, the electronic expansion valve, to operate properly. In addition, the level of subcooling has a direct relationship with chiller capacity. A reduced level of subcooling results in a shift of state 3 (in FIG. 6B) to the right and a corresponding shift of state 4 to the right, thereby reducing the heat removal capacity of the evaporator (Q1).
As the chiller's refrigerant charge increases, the accumulation of refrigerant stored in the condenser on the high-pressure side of the system also increases. An increase in the amount of refrigerant in the condenser also occurs as the load on the chiller decreases due to less refrigerant flow through the evaporator, which results in increased storage in the condenser. A flooded condenser causes an increase in the amount of sensible heat transfer area used for subcooling and a corresponding decrease in the surface area used for latent or isothermal heat transfer associated with condensing. Therefore, increasing refrigerant charge level and decreasing chiller load both result in increased subcooling temperatures and condensing temperatures.
Increased outdoor air temperatures have an opposite effect on the operation of the condenser. As the outdoor air temperature increases, more condenser surface area is used for latent or isothermal heat transfer associated with condensing and a corresponding decrease in sensible heat transfer area used for subcooling. Therefore, increases in outdoor air temperature result in decreased subcooling temperatures and increased condensing temperatures.
Referring to FIG. 6B, an increase in subcooling drives state 3 to the left, while an increase in condensing temperature shifts the curve connecting states 2 and 3 upward. High condensing temperatures can ultimately lead to compressor motor overload and increased compressor power consumption or lowered efficiency. As subcooling increases, heat is added to the evaporator, resulting in an upward shift of the curve connecting states 4 and 1. As the evaporating temperature increases, the specific volume of the refrigerant entering the compressor also increases, resulting in increased power input to the compressor. Therefore, increased levels of refrigerant charge and decreased chiller load conditions result in increased subcooling, which leads to increased compressor power input.
Control of the electronic expansion valve is based on a sensor located within the compressor's inlet where it measures superheat level. Superheat level is represented by the slight increase in temperature after the refrigerant leaves the saturation curve, as shown at state 1 in FIG. 6B. Vaporized refrigerant leaves the chiller's evaporator and enters the compressor as a superheated vapor with a recommended setpoint (2.2 C) superheat to avoid premature failure from droplet pitting and erosion.
As discussed previously, an increase in outdoor air temperature causes an increase in discharge pressure, which, in turn, causes the compressor's suction pressure to increase. The curves connecting states 2 and 3 and states 4 and 1 on FIG. 6B 3 both shift upward due to increases in outdoor air temperature. An upward shift in curves 4 through 1 or an increase in refrigerant evaporating temperature results in a decrease in the evaporating approach temperature. As the approach temperature decreases, the mass flow rate through the evaporator must increase in order to remove the proper amount of heat from the chilled water loop. Therefore, increasing outdoor air temperatures cause evaporating pressure to increase, which leads to increased refrigerant mass flow rate through the evaporator. The combined effect of higher refrigerant mass flow rate through the evaporator and reduced approach temperature causes a decrease in superheat temperatures. Therefore, an inverse relationship exists between outdoor air temperature and superheat temperatures.
With decreasing refrigerant charge, the curve connecting states 2 and 3 in FIG. 6B shifts downward and the subcooling level decreases or state 3 on the T-s diagram in FIG. 6B moves to the right. Bubbles begin to appear in the liquid line leading to the expansion device due to an increased amount of gaseous refrigerant leaving the condenser. Without the proper amount of subcooling in the refrigerant entering the expansion device (state 3 in FIG. 6B), the device does not operate optimally. In addition, a decrease in refrigerant charge causes a decrease in the amount of liquid refrigerant that flows into the evaporator and a subsequent decrease in capacity and increase in superheat and suction pressure. Thus, an inverse relationship exists between refrigerant charge level and superheat temperature.
Under extreme refrigerant undercharge conditions (below −20% charge), refrigerant undercharge causes an increase in suction pressure. In general, the average suction pressure increases with increasing refrigerant charge during all charge levels above −20%. Refrigerant charge level is a significant variable in determining both superheat temperature and suction pressure.
U.S. Pat. Nos. 4,437,322; 4,858,681; 5,653,282; 4,539,940; 4,972,805; 4,382,467; 4,365,487; 5,479,783; 4,244,749; 4,750,547; 4,645,542; 5,031,410; 5,692,381; 4,071,078; 4,033,407; 5,190,664; and 4,747,449 relate to heat exchangers and the like.
There are a number of known methods and apparatus for separating refrigerants, including U.S. Pat. Nos. 2,951,349; 4,939,905; 5,089,033; 5,110,364; 5,199,962; 5,200,431; 5,205,843; 5,269,155; 5,347,822; 5,374,300; 5,425,242; 5,444,171; 5,446,216; 5,456,841; 5,470,442; 5,534,151; and 5,749,245. In addition, there are a number of known refrigerant recovery systems, including U.S. Pat. Nos. 5,032,148; 5,044,166; 5,167,126; 5,176,008; 5,189,889; 5,195,333; 5,205,843; 5,222,369; 5,226,300; 5,231,980; 5,243,831; 5,245,840; 5,263,331; 5,272,882; 5,277,032; 5,313,808; 5,327,735; 5,347,822; 5,353,603; 5,359,859; 5,363,662; 5,371,019; 5,379,607; 5,390,503; 5,442,930; 5,456,841; 5,470,442; 5,497,627; 5,502,974; 5,514,595; and 5,934,091. Also known are refrigerant property analyzing systems, as shown in U.S. Pat. Nos. 5,371,019; 5,469,714; and 5,514,595.